Turbine steam admission controls



March 22, 1966 c. STROHMEYER, JR

TURBINE STEAM ADMISSION CONTROLS Filed Feb. 11, 1965 4 Sheets-Sheet 1 Fig.1.

INVENTOR.

Charles Strohmeyer Jr hi5 ATTORNEY 4 Sheets-Sheet 2 NOZZLE CHEST E' INVENTOR Charles Strohmoyeqd:

I00 BY hisATTORNEY March 22, 196 c. STROHMEYER, JR

TURBINE STEAM ADMISSION CONTROLS Filed Feb. 11, 1965 NOZZLE CHESTS A'& A"

LOAD, OF RATING March 22, 1966 c, STRQHMEYER, JR 3,241,322

TURBINE STEAM ADMISSION CONTROLS Filed Feb. 11, 1963 4 Sheets-Sheet 4 Fig.6. 5 f/Al '65 l I25 A @lllll C harles Sfrohmeyer, Jr

his ATTORNEY lI/I/I United States Patent 3,241,322 TURBINE STEAM ADMISSION CONTROLS Charles Strohmeyer, Jr., Wyomissing, Pa., assignor to Gilbert Associates, Inc., Reading, Pa. Filed Feb. 11, 1963, Ser. No. 257,674 6 Claims. (Cl. 60-73) This invention relates to methods, devices and systems for improving the control of admission steam to a steam turbine which is the prime mover for a steam-electric generating plant and includes coordinated pressure and fiow control of the steam from the steam generator to the steam turbine.

The invention also relates to a steam electric generating plant having a steam turbine and steam generator of the reheat type.

An object of the present invention is to provide a means for controlling steam temperatures Within the turbine which will increase turbine cycle efficiency, particularly with respect to the reheat portion of the cycle.

A further specific object of the invention is to control steam pressure drop across the turbine first expansion stage below a preselected value in a way which effectively reduces first stage blade stress for high pressure applications and where steam is admitted to the said first expansion stage in stepped valved increments, each valve increment furnishing steam to a partial arc of the first stage blades.

A still further specific object of the invention is to provide a means for coordinating operation and control of the turbine governor valves with a pressure reducing valve located internally within the steam generator circuits.

Other objects and advantages of the present invention will become more apparent from a study of the following description taken with the accompanying drawings where- FIG. 1 is a schematic diagram of the steam generator circuits, turbine steam admission components and turbine cycle embodying the present invention;

FIGS. 2a and 2b are transverse and longitudinal crosssections, respectively, of a typical high pressure turbine in the zone of the first expansion stage;

FIG. 3 is a plot of steam admission pressures to the high pressure tubine vs. load;

FIG. 4 is a plot of steam enthalpy and temperature vs. pressure;

FIG. 5 is a diagram of a control system for controlling inlet steam pressure to the high pressure turbine;

FIG. 6 is a diagram of a speed control system for the turbine governor valves;

FIG. 6a shows the gearing arrangement between the pivot positioner and turbine governor valves; and,

FIG. 7 is a modification of FIG. 5 control system for controlling inlet steam pressure to the high pressure turbine.

General description of present invention A simple flow diagram of a steam electric generating plant covered by this invention is shown in FIG. 1. Steam generator 1 indicated in dash and dot outline has preheating section 2, generating section 3, primary superheater 4, and secondary superheater 5, which sections are serially connected by fluid conduit means. Throttling valve 6 is shown in the conduit between generating section 3 and superheater 4. Valve 6 could be alternatively located between superheaters 4 and 5. The criterion for this invention with respect to valve 6 is that valve 6 be located in the steam generator circuits upstream of at least a portion of the superheater.

Conduit 7 conduits steam from superheater 5 through conduits 7a and 7b to turbine stop valves 8a and 8b. Stop valves 8a and 8b are provided with hydraulic servo- 3,241,322 Patented Mar. 22, 1966 motor operators 9a and 9b. Stop valves 8a and 8b discharge to governor valve chests 10a and 10b. Each governor valve chest has four governor valves designated as A, B, C, D or E. All of the same letter valves are operated together in parallel. The valves A are opened first followed by B, C, D and E in sequential order. Valve E is closed first followed by D, C, B and A in sequential order.

Each governor valve is provided with an operating shaft 11, which is connected to beam 12 by means of a pin joint at 13. The pin support in beam 12 is slotted to produce the required valve movements. The beam 12 is supported at one end by fixed strut 14 through a pin joint 15. The beam 12 pivots about 15. At the other end, beam 12 is attached through pin joint 16 to piston shaft 17 which is actuated by servomoter power mechanism 18a or 1817. Movement of shaft 17 up and down raises and lowers beam 12 as it rotates about pivot 15. This, in turn, causes valves A, B, C, D and E to open in sequential order. Servomotors 18a and 18b are operated in parallel. Shafts 11 are provided with collar 19 upon which spring 20 exerts a downward force to oppose the opening of the governor valves. The springs 24 are retained by brackets 21 which are attached to the governor valve chests 10a and 10b.

Governor valves A discharge through conduits 22, 23, 24 and 25. Conduits 22 and 23 discharge to turbine first expansion stage nozzle area A. Conduits 24 and 25 discharge to turbine first expansion stage nozzle area A. Conduits 26, 2'7, 28 and 29 discharge to turbine first expansion stage nozzle areas B, C, D and E, respectively.

The high pressure first expansion stage nozzle areas A, A, B, C, D and E are located in high pressure turbine 30. FIG. 2a shows a transverse cross-section of the high pressure turbine through the first expansion stage nozzle chest. FIG. 2b shows a longitudinal cross-section of the high pressure turbine through the same area. The relative locations of FIGS. 2a and 2b are shown in FIG. 1.

In FIG. 1 the steam exhausts from high pressure turbine 30 through conduit 31 to series reheaters 32 and 33 which are located in steam generator 1, and which raise the temperature of the steam. Reheater 33 discharges through conduit 34, through stop valve 35 and interceptor valve 36 to intermediate pressure reheat turbine 37. Turbine 37 is of the double flow type and exhausts through conduits 38 to two double flow low pressure turbines 39. Turbines 39 exhaust through conduits 40 to condenser 41. Cooling water is circulated through con duits 42. Turbines 30, 37 and 39 are connected together by shafting means 51 to electric generator 52.

Condensed condensate is collected in hotwell 43 and is drawn through conduit 44 to the suction of pump 45 driven by motor 46 which raises the fluid to the first, working pressure of the cycle. Pump 4-5 discharges through conduit 47, through low pressure heater system 48 to high pressure feedwater pump 49 driven by motor 50.

Feed pump 49 discharges through conduit 53 to high pressure heater system 54, through conduit 55 to the steam generator 1 preheater 2.

FIGS. 2a and 2b show tranverse and longitudinal cross sections of high pressure turbine 60, respectively. The high pressure turbine has an outer casing 56 and inner casing 57. The casings are split on the horizontal center lines and are bolted together.

In FIGS. 2a and 2b the high pressure first expansion stage nozzles areas are shown as nozzle chests A, A, B, C, D and E. The nozzle chests and connecting conduits are separated one from the other from the inlet of the first stage stationary blades 58 back to the governor valve seats in chests 10a and 10b as shown in FIG. 1.

In FIGS. 2a and 2b the stationary first stage blades 58 are mounted in the nozzle chests A, A, B, C, D and E which are guided and held in place by means of inner casing 57. The first expansion stage is of the impulse type. There are two rows of rotating blades 59 mounted on turbine rotor 60stationary blades 58 reduce the steam pressure and form the stage inlet nozzles-and stationary blades 61 reverse the steam flow in this velocity compound stage. Blades 61 are held by ring 6-2 which is fitted and supported in the inner cylinder 57. Around the ring 62 are open areas 63 between the ring 62 and the inner casing 57 through which steam may pass. For example, flow passes through conduit 2223 to nozzle chest A, through blades 58, the first row of blades 59, blades 61, the second row of blades 59 to area 64, through the areas 63 to area 65. In flowing from area 64 to area 65, the steam flows around the nozzles chests A, A", B, C, D and E, cooling them.

From area 65 the steam passes through the subsequent downstream steam expansion stages 66, 67, 68, 69, etc. Each subsequent stage consists of a row of stationary and rotating blades. The steam finally exhausts from high pressure turbine 30 as shown in FIG. 1.

The 'FIG. 3 vertical scale shows steam pressure as a percent of turbine inlet design operating steam pressure for steam in nozzle chests A, A, B, C, D and E and, at the first stage exhaust in area 65 for various modes of operation and for loads ranging from 0 to 100 percent on the horizontal scale.

In conventional past practice the steam generator has been operated to hold the steam pressure at the turbine inlet constant. Pressure at the discharge of the first expansion stage is represented by the line FG in FIG. 3. This line is approximately straight over the load range and is dependent upon fiow conditions from the second stage inlet to the condenser. There is no controlled throttling of flow between the two points.

If the inlet steam pressure to the turbine 30 is held constant at design level, then the governor valves must reduce the steam pressure before the nozzle chests as load is reduced. Governor valve E is the first to close. As it closes, pressure in the nozzle chest E approaches first stage exhaust pressure. Line H-J represents the pressure in nozzle chest E throughout the control range of valve E. The lines K-L, M-N, P-Q represent equivalent pressures in nozzle chests D, C and B as a result of action of governor valves D, C and B closing sequentially. As load is further decreased, governor valves A are closed and pressures in nozzle chests A and A follow the line F-U-R. Pressure drop through valves A at this time is from line T-R vertically down to line FU-R.

Where design steam pressure to high pressure turbine 30 is, for example, 3500 p.s.i.g., the first valve point (R) is near the full load rating. Step valve control range (R to J) is limited. Therefore, partial load efficiencies below load R have greater significance as inlet steam pressures are increased over past practice. Also, it will be noted that the greatest pressure drop through the first stage nozzles occurs at load R, the pressure drop being represented by the vertical distance from R to P. Mass flow through blades 58, 59 and 61 in the A and A nozzle chamber segments is greatest at this time. Also, there is no flow through the other nozzle chamber segments. This results in considerable shock to the first stage rotating blades 59, and stationary blades 61.

First consider partial load efficiencies below point R. Referring to FIG. 4, steam enthalpy, B.t.u./lb., is indicated on the left side vertical scale. Steam pressure is indicated on the bottom horizontal scale and the sloping constant steam temperature lines are indicated on the right side vertical scale. For the case where the design inlet steam temperature to high pressure turbine 30 is 3500 p.s.i.a. and 1000 F. and at partial load the nozzle chest pressure is as shown at S in FIG. 3, and the inlet steam is maintained at design pressure and temperature throughout the load range, there is a 50 percent pressure drop through governor valves A.

From FIG. 4 it can be seen that when 3500 p.s.i.a., 1000F steam is reduced in pressure adiabatically to 1750 p.s.i.a. at constant enthalpy through a valve, after expansion is complete, the temperature is changed to approximately 920 F. This temperature change occurs without any work having been extracted from the steam and is associated solely with the physical properties of the steam. Thus, the steam entering the first stage nozzle chests A and A at this time is below design temperature of 1000 F.

There is an established ratio of absorption in steam generator 1 between the high pressure primary circuit and the reheat circuit for each load condition. This ratio usually may be varied slightly through control means. The temperature level in conduit 31 influences maximum temperature possible in conduit 34. Thus, as steam temperature in the nozzle chests falls below design, high reheat steam temperature capability also decreases. When hot reheat steam temperature decreases substantially below design values, turbine efficiency is impaired.

From FIG. 4 it can be seen that for load S, if the steam were delivered to the nozzle chests at 1750 p.s.i.a. and 1000 F., the enthalpy of the steam would be increased approximately 55 B.t.u./ lb. This enthalpy increase would be reflected throughout the flow path of the turbine including elements 30, 37 and 39. Also, the increased enthalpy or heat content of the high pressure steam would require addition of more B.t.u.s/lb. to the flow between the inlet of preheater 2 and the outlet of superheater 5. This in turn makes more heat available to rehcaters 32 and 33 per pound of flow as a result of the related absorption characteristics between the two circuits for each load condition. Thus, by minimizing temperature drop in the nozzle chests A and A below load R, turbine cycle eificiencies can be increased particularly in the reheat portion of the cycle.

FIG. 4 shows that if the pressure of the steam in conduits 7a and 7b (FIG. 1) is maintained at 3500 p.s.i.a. for load S (FIG. 3), the temperature in conduits 7a and 7b would have to be approximately 1080 F. to produce 1000 F. in nozzle chests A and A". This would be impractical as the superheater 5 would have to be designed for excessively high temperatures which would only exist during partial load conditions.

As an alternate, it has been shown in my copending patent application, Serial No. 42,194, filed June 7, 1960, that superheater outlet steam enthalpy can be increased at partial loads without exceeding full load rated steam temperature by reducing steam pressure upstream of at least a portion of the superheater. In FIG. 1 of this application, throttling valve 6 is capable of reducing steam pressure down to the limit required to produce desired steam pressure in the nozzle chests.

A novel feature of this invention relates to the operating method and control means to optimize turbine cycle efiiciency, particularly in the reheat portion of the cycle and in conjunction with regulation of steam pressure downstream of valve 6.

Referring to FIG. 3, the conventional past method of plant operation has been to control steam pressure in conduit 7 (FIG. 1) along the line T-R J. Turbine cycle efficiency can be improved by increasing the enthalpy of the steam in conduit 7 within design temperature limits at the outlet of the superheater and by reducing pressure in conduit 7 below the line TR down to the limit re quired to produce the F, S, R line pressure in the nozzle chests A and A for various partial loads below load R through the use of control means described in this invention.

Next consider maximum pressure drop across the first stage nozzles as represented by the vertical line R-P in FIG. 3. The pressure drop across the first stage nozzles for past conventional methods of plant operation is the vertical distance between the lines F-U-R-J and F-G. If, however, nozzle chest pressure were limited to the line between points U and V, pressure drop across the first stage would be limited to a maximum preselected value. This in turn reduces maximum blade stress and shock between loads U and V. For this method of operation the governor valves A would be fully open at U. Line W-X would be the pressure in nozzle chest B for the control range of governor valve B. Lines Y-Z and AAL would be the pressure in nozzle chests C and D for the control range of governor valves C and D respectively.

Another novel feature of this invention relates to the operating method and control means to limit pressure drop across the first stage to a maximum preselected value which is below that which would result from design steam pressure in the operating nozzle chests when the first step governor valves (valves A) are first in the fully open position when increasing load as at position R.

It is understood that operation of valves A, B, C, D and E can be overlapping as a substitute for the purely sequential operation shown in FIG. 3

There are many possible ways of accomplishing the above objectives. The following descriptions of control means are illustrative of this invention. For purposes of description simplicity, pressure drop through the stop valves 8a, 8b, valve chest a, 10b, governor valves in the wide open position and connecting conduits between valve chests 10a, 10b and the nozzle chests is ignored and considered zero. In actual application, allowance is made for these factors.

Inlet steam pressure to the turbine from conduit 7 can be controlled along the FIGURE 3 lines BB-UV'J or F-U-V-J for both nozzle chest steam temperature control and to limit maximum first stage pressure drop. Inlet steam pressure can be controlled along the line F- U-R-I if nozzle chest steam temperature control only is desired. Inlet steam pressure can be controlled at some other higher level as along line CC-DD for partial advantage. 7

Inlet steam pressure in conduit 7 can be controlled by raising and lowering the pressure in the overall circuit from the feedwater pump 49 discharge (FIG. 1) to the turbine 30 steam inlet. However, many types of high pressure steam generators, especially in the supercritical pressure range, cannot reduce pressure substantially in their circuits as load is decreased. This invention is di rected particularly toward this type of unit and which includes valve 6 in the steam generator circuits. In the case where it is possible to reduce steam generator operating pressure with load, capability for pressure change is generally slow. Therefore, control of steam pressure in conduit 7 by the use of valve 6 is very practical and convenient during a rapid pressure transition period for this type of unit. Turbine speed governor control principles for this invention will apply in the latter case.

It has been stated that increased enthalpy of primary steam to turbine 30 will increase temperature and enthalpy of steam to reheat turbine 37 and thus improve reheat turbine efiiciency. It should be understood that for any given combined power output from turbines 30, 37 and 39, the higher primary steam enthalpy and hotter reheat steam temperature will require more heat input to the fluid in the steam generator per pound of fluid flow. However, the total steam flow for the said given power. output will be lower where the steam heat content per pound is higher. As a result of the improved turbine eificiency, the total heat input to the steam generator willbe less as temperature in the turbine nozzle chests is increased as described above for this partial load condition.

Firing means for steam generator 1 are shown in FIG. 1. Fuel is flowed to burner 195 through conduit 196. Valve 198 controls the rate of flow and is actuated by power operator 197. Air flow enters burner 195 through conduit 199. Damper 200 controls the rate of air flow and is actuated by power operator 201. Steam temperature at point 202 is sensed by a thermocouple and conduit leads 203 convey the potential generated to measuring instruments in the combustion control system 204. Combustion control system 204 is standard and is designed to maintain a programmed or constant steam pressure and temperature at points 207 and 202 respectively. System 204 may be manual or automatic. 'It may be of the analog or digital computor type. Combustion control system 204 sends control signals to valve operators 197 and 201 through conduits 205 and 206 respectively.

Thus, steam enthalpy can be controlled in conduit 7 by control of steam pressure downstream of valve 6, temperature at point 202 and firing rate in burner 195. High pressure once-through type steam generators can be designed so that burner 195 can be fired at a rate sufiicient to substantially increase steam enthalpy at point 202 as pressure is reduced downstream of valve 6 and within the designed temperature limit of the final super heater outlet.

FIG. 5 shows one means for control of valve 6. The objective of the control system is to control pressure in conduit 7 (FIG. 1) with load in a manner consistent with the shape of one of lines BBUVI or CC-DD-J as shown in FIG. 3 and as is desired for the mode of operation to be employed. Load may be measured from any one of several indexes as generator 52 electrical output, turbinefirst stage exhaust steam or steam flow. The latter is more diflicult to measure because of fluctuating pressures downstream of valve 6.

In FIG. 5, turbine first stage exhaust steam pressure is. used as a measure of turbine generator load. First stage exhaust pressure is measured in area 64 as shown in FIG. 2b. Area 64 pressure is conducted through conduit 70 to pressure transmitter 71 shown in FIG. 5. Transmitter 71 converts the high pressure of the steam to a low pressure (i.e. 3-27 p.s.i.g.) pneumatic control air signal in conduit 72. Air pressure changes in conduit 72 are directly proportional to steam pressure changes in conduit 70. Conduit 72 feeds to function generator 73 which characterizes the output signal in conduit 74 with the desired pressure line from FIG. 3 as mentioned above. Therefore, the air pressure in conduit 74 is proportional to and is a measure of the desired steam pressure in conduit 7.

The function generator 73 may be a computing relay as manufactured by the Bailey Meter Company, Cleveland, Ohio and as shown in their Produce Specification P.9913, copyright1962. The formula for the relay is )f( A equals control air pressure in conduit 72, D equals control air pressure in conduit 74, X and Y are constants, f is a function generated by a cam actuated by the changing control air pressure A. The shape of the cam characterizes the output signal D. Therefore, to characterize FIG. 3 pressure line BB-U-VJ and where 27 p.s.i.g. air pressure in conduit 74 equals design steam pressure in conduit 7 (FIG. 1) or the wide open position of valve 6 and where 3 p.s.i.g. air pressure in conduit 74 equals zero steam pressure in conduit 7 or the closed position of valve 6, function f is calibrated so that the output signal D in conduit 74 is maintained at 27' p.s.i.g. from J to V (FIG. 3) and is gradually reduced as first stage pressure decreases to proportionately follow line V-U-BB The lines BB-V or CCDD can be characterized as a straight sloping line by the simple formula i (A X) Percent Proportional Band Where the proportional band setting or gain is held constant and the constant Y produces the maximum output control signal (i.e. 27 p.s.i.g.) in the load range from V to J or from DD to 1, respectively.

The firing rate and feedwater flow controls for steam generator 1 shown in FIG. 1 would normally be associated with fluid pressure control upstream of valve 6, where valve 6 was used to vary downstream pressure throughout a portion of the load range. The firing rate and feedwater flow controls will be coordinated with operation of valve 6 to avoid wasting pumping power, especially in the range of designed steam flow. Minimum pressure drop across valve 6 is desired at this time. One method of integrating feedwater flow control with firing rate control by automatic means is shown in FIG. 3 of my copending application Serial No. 251,358, filed January 14, 1963, now Patent No. 3,186,175. Provision for automatic regulation of feedwater flow control is not shown in FIG. 1.

Conduit 74 in FIG. 5 feeds to proportional plus reset action computing relay 75. Relay 75 may be as manufactured by the Bailey Meter Co., type AS2300 and as shown in FIG. 6 on page 3 of their product specification P992, copyright 1957 by Bailey Meter Co., form CP99-2B printed July 1960. The signal in 74 feeds to the A chamber and acts as a set point in relay 75.

Steam pressure at the inlet to turbine 30 is measured at point 76 in conduit 7 shown in FIG. 1. Pressure from point 76 is conducted through conduit 77 to pressure transmitter 78 shown in FIG. 5 which is similar to transmitter 71. Pressure transmitter 78 converts the steam pressure in conduit 77 to a low pressure pneumatic control air signal (i.e. -3 to 27 p.s.i.g.) in conduit 79. Air pressure changes in conduit 79 are proportional to steam pressure changes in conduit 77. Conduit 79 feeds to the B connection of relay 75.

Relay 75 functions so that when the proportional control air pressure (actual steam pressure) in conduit 79 is below the characterized set point control air pressure (required steam pressure) in conduit 74, the control air output signal D in conduit 80 gradually increases until the pressures in conduits 79 and 74 are balanced. Thus, actual steam pressure is compared with and corrected to required steam pressure in main steam conduit 7 (FIG. 1). Conversely when the actual steam pressure is above required steam pressure, output signal D control air pressure in conduit 80 gradually decreases until required and actual pressures are balanced. The formula for relay 75 is A is pressure in conduit 74, B is pressure in conduit 79. X and Y are constants, G=gain or 100% proportional band and Ti=integrating time constant, dt=increment of time.

Conduit 80 feeds to valve positioner 81. Valve 6 is equipped with hydraulic servomoter 82, which has piston 83 and connecting shaft 84 for operating valve 6. Hydraulic working fluid is fed to servomotor 82 through relay 85, which selectively controls fluid flow to each side of piston 83, through conduits 86 and 87. Hydraulic working fluid supply relay 85 is from reservoir 88, through conduit 89 to pump 90 through conduit 91 to relay 85. The relay 85 drains through conduits 92, 92a and 92b to reservoir 88. A relief valve 93 is located in the discharge conduit 91 to prevent overpressure from pump 90. Accumulator 94 provides volume storage for the hydraulic working fluid and is equipped with a flexible sealed sack 94a. Hydraulic working fluid is contained in the space 94b and the space 94c is filled with a high pressure gas. The relay valve 95 is connected to shaft 96 which is actuated by diaphragm mechanism 97.

Air pressure in chamber 98 presses on diaphragm 99 and opposes spring 100. When te air pressure on diaphragm 99 is balanced with the spring 100, the relay valve 95 is in neutral position and appropriate pressures are maintained on piston 83 to maintain valve 6 in its required position to maintain superheater pressure to set point as established in conduit 74.

Valve positioner 81 is connected to valve connecting shaft 84 through mechanical link 101. Link 101 actuates a cam (not shown) in positioner 81, the position of which is indicative of the position of shaft 84 and valve 6. Control signal changes in conduit are reflected in conduit 102, which in turn, upsets the balanced position of relay valve 95. This causes the pressures to change to both sides of the piston which moves the piston and shaft 84.

The cam is positioner 81 is shaped so that when shaft 84 and valve 6 position move to the new position consistent with the changed pressure in conduit 80, the cam movement and its effect on pilot valve action in positioner 81 will cause a balanced condition between the air pressure in chamber 98 and the spring 100. This in turn restores relay valve to the neutral position.

The cam in positioner 81 can be shaped so that valve 6 opening will produce downstream pressure changes which are linear to pressure changes in conduit 80. FIG. 5 can be constructed from standard type commercially available control equipment components.

Steam pressure variations at the inlet to the high pressure turbine 30 will have an effect upon the turbine speed governor controls. The turbine speed governor for this type of application usually has about 5 percent of rated speed control range from the fully open to fully closed governor valve position. The turbine governor speed control characteristics should be coordinated with control of valve 6. Where inlet steam pressure in conduit 7 (FIG. 1) is reduced with load, independently of a direct tie with the turbine governor, the turbine speed governor must be able to continuously close the governing valves with an increase in turbine speed and open the governing valves with a decrease in turbine speed, otherwise the FIG. 5 system will be inoperable. For example, inlet steam pressure in conduit 7 could not be controlled from load to follow the line FUR in FIG. 3. In this case governor valve and valve 6 control actions would be overlapping. The valve 6 action would prevent a change in governor valve position with a change in load, yet valve 6 action is dependent upon a change in load initiated from a change in governor valve position.

Where it is desirable to control inlet stream pressure along a line F-S-U or F-U-R as shown in FIG. 3, it will be necessary to transfer turbine speed govering control from the governing valves to valve 6, locking the governor valves in a fixed position at the point of transfer.

As inlet steam pressure in conduit 7 is reduced, it may be desirable to automatically change the proportional setting of the turbine governor speed range in a characterized manner so that the turbine generator may be operated satisfactorily in parallel with other electric generating units. For example, if the inlet steam pressure as well as the opening of the governing valves is reduced with an increase of speed, the load drop will be considerably greater than would occur if the inlet steam pressure were held constant and the governing valves only were closed from the change in speed. Increasing the proportional range of the turbine speed governor as inlet steam pressure is decreased can be employed to parallel governor action for the two inlet steam pressure control conditions (variable and constant).

FIG. 6 shows a hydraulic type of turbine governor speed control and illustrates adaptations to accomplish the objectives of this invention. The turbine shaft 51 (FIG. 1) speed is measured by the speed governor. This speed measurement is then used to control opening and closing of governor valves A, B, C, D and E. Governor valve openings are controlled by hydraulic servomotors 18a and 18b.

In FIG. 6, the turbine shaft section 51 shown is at the end opposite to the generator 52 (FIG. 1) and the shaft has a governing speed pump impeller 183 mounted therein. Flow is forced through the impeller 183 in reverse. High pressure oil is supplied to impeller 183 through conduit 104. Orifice 105 controls the quantity of oil flow. Oil supply enters ring 106 and flows back through impeller 163 into cavity 107 and discharges through drain conduit 168. Thus, the pressure in ring 106 is representative of the speed of shaft 51. The oil pressure varies as the square of the speed.

Ring 106 supplies oil to and controls pressure in conduit 109 which feeds to chamber 110. Oil in chamber 118 surrounds bellows 111. Bellows 111 actuates rod 112 which is connected to beam 113. Beam 113 pivots about point 114, alignment and position of which is maintained by gears 115 and 116. Beam 113 has a needle point extension 117 which controls the opening of cup valve 118. The beam 113 is spring loaded on the top side by spring 119. The compression of spring 119 is regulated by screw 120. Screw 120 may be turned by handwheel 121 and shaft 122 or may be operated by motor 123 through worm gear 124 and companion gear 125. Gear 125 is provided with a slip clutch (not shown) to permit operation of screw 120 through handwheel 121.

Pressure in chamber 110 acting on bellows 111 is trans mitted through shaft 112 to beam 113. This pressure opposes the pressure on beam 113 from spring 119. An increase in chamber 110 pressure also relieves the pressure from needle point 117 on cup valve 118. Pressurized oil enters conduit 126 and flow quantity is controlled by orifice 127 before it enters chamber 128. Thus oil pressure in chamber 128 is controlled by overflow through cup valve 118. Overflow is controlled by the downward pressure from needle point 117. Valve 129 is a relief valve. Spring 130 limits maximum pressure in chamber 128.

The base range of speed control is established by the compression spring 119. Thus, where controlled speed is 3600 r.p.m., the desired loading of the machine from minimum to full load is established by speed changer handwheel 121 or speed changer motor 123. When speed (or frequency of electrical output as 60 cycles A.C.) varies, the turbine will react to the speed governor and increase or decrease load accordingly above and below the base setting of the speed changer.

.The oil pressure in chamber 128 is transmitted through conduit 131 to the hydraulic servomotors 18a and 18b. Conduit 132 discharges to servomotor 18a and conduit 133 discharges to servornotor 181: (not shown) which is the same as and is operated in parallel with servomotor 18a. Conduit 134 is normally closed off by valve means (not shown). In times of emergency oil flow can be dumped through conduit 134 to quickly close servomotors 18a and 1812.

'For the following discussion, assume valve 135 open and valves 136 and 137 closed. Flow from conduit 132 passes to chamber 138 where it pushes against relay piston 139. Shaft 140 with built in cup valve 140a is connected to piston 139 and presses against relay plunger 141.

Pressurized oil enters conduit 142 and flows around ring 143 and enters ports 144 in relay bushing 145 to chamber 146 in between relay plunger 141 and relay bushing 145. Flow from chamber 146 passes through conduit 147 and orifice 148 to chamber 149. Spring 150 pushes relay plunger 141 up against cup valve 140a. Increased pressure in chamber 149 causes relay plunger 141 to separate from cup valve 140a. Flow through cup valve 140a and conduit 150 to drain 151 causes relay plunger 141 to follow relay piston 139.

'Increased pressure in chamber 138 moves piston 139 and relay plunger 141 downwardly so that fluid from space 146 can flow through port 152 in relay bushing 145. Flow through port 152 discharges to space 153 under piston 154 moving piston 154 and tappered connecting shaft 155 upward. Shaft 155 connects to drive shaft 17 which actuates lever arm 12 through pin joint 16 opening the governor valves A, B, C, D and E. As the tappered shaft rises, follower arm 158 rotates in a clockwise direction around pin 156 in bracket 157. This causes relay bushing to move downward closing the gap between space 146 and port 152. The mechanism is so constructed that for each control oil pressure in chamber 138 there is a specific position for piston 154.

Decreased pressure in chamber 138 causes 139 and 141 to rise permitting oil in space 153 to pass through port 152 to space 159 from whence it flows to conduit and to drain 151. As the tappered shaft drops as a result of spring pressure on the individual governor valves A, B, C, D and E, the follower arm 158 rotates counterclockwise which raises the relay bushing 145 and closes the gap between spaces 152 and 159 stopping the flow of oil from space 153 when piston 154 reaches the preselected position for the changed pressure in chamber 138.

Tension on spring 160 may be adjusted by turning nut 161a on stud bolt 161. This permits positioning of piston 154 with relay piston 139.

The above description of the turbine hydraulic speed governor covers the basic conventional features of the system. a

The proportional range of the turbine speed governor can be changed as shown in FIG. 6. The position of pivot point 114 is controlled by floating gears 115 and 116. Gears 115 and 116 move in rack teeth cut in the top and bottom of beam 113 and pivot positioner 162. As positioner 162 moves to the left, gear 115 rotates counter-clockwise, gear 116 rotates clockwise and point 114 moves to the left. Beam 113 remains :fixed as movement of gears 115 and 116 is controlled by fixed racks 163 on each side of beam 113 as shown in FIG. 6a. Movement of positioner 162 to the left increases the oil pressure in chamber 128 for any given pressure in cham-- ber 1110 and thus positions the governor valves in a more open position. Movement of positioner 162 to the right decreases the oil pressure in chamber 128 and thus positions the governor valves in a more closed position.

Positioner 162 movement can be coordinated with the control system shown in FIG. 5 so that speed control and load change characteristics for variable inlet steam pressure will be the same as they are for constant inlet steam pressure. The speed control load change characteristics can also be characterized to any other desired function with load. Control of positioner 162 in FIG. 6 is from load signal as indicated by turbine first stage exhaust steam pressure transmited through conduit 70 (FIG. 2b). Pressure transmitter 164 and function generator 166 are similar to items 71 and 73 for FIGURE 5 respectively and function in the same manner. Thus, function generator 166 characterizes the movement of positioner 162 with load. Items 164 and 166 are connected through conduit 165. Control air output of function generator 166 is transmitted through conduit 167 to the diaphragm conroller 169 of piston operator 168.

Increased control air pressure in 169 through diaphragm 170 compresses spring 171 and moves shaft 172 and relay valve 173 to the right. This permits power air in conduit 174 to enter chamber 175 through conduit 176 which moves piston 177 to the right, also chamber 178 is vented through conduit 179. As the piston 177 and shaft 188 move to the right, cam 181 causes follow up lever 182 to rotate clockwise around fixed pivot 183. This increases the compression on spring 171 to return shaft 172 and relay valve 173 to the neutral position. Thus, for each control air pressure in conduit 167 there is a set position for piston 177, shaft and pivot positioner 162.

Decreased control air pressure in 169 moves shaft 172 and relay valve 173 to the left. Power air enters chamber 178 and piston 177, shaft 180 and pivot positioner 162 move to the left. Movement of cam 181 to the left rotates follow-up lever 182 counter-clockwise which reduces compression on spring 171 and returns relay valve 173 to the neutral position when shaft 181} has reached a predetermined position.

Thus, the control system of FIGURE in conjunction with control of pivot positioner 162 (proportional control) can be use together for the case where high pressure tur bine inlet steam pressure is controlled along lines BBUVJ or CCDDJ as shown in FIG. 3.

In the case where high pressure turbine inlet steam pressure is controlled along line FURJ as shown in FIG. 3, a different technique is required. For this case refer to FIG. 6. Pivot point 114 is flexible but fixed. Positioner 162 and associated controls are not required. The speed governor is operated in a conventional manner between FIGURE 3 points R and J. Valve 135 is closed and valves 136 and 137 are open. Pressurized oil enters through conduit 184 and pressure reducing valve 185 maintains a minimum control oil pressure in conduit 186. Conduit 186 feds to chamber 138. The minimum pressure is set so that governor valves A will always remain open except in case of emergency when they are rapidly closed. Check valves 187 and 188 prevent reverse flow in conduits 184 and 186. Thus, pressure in conduit 131 may fall below pressure in conduit 186 for loads below point R in FIG. 3. Oil leakage from space 138 through piston 139 and shaft 140 clearances causes pressure in space 138 to follow pressure in conduit 131 above minimum values. Flow from conduit 133 in this case passes to the controls for valve 6 as shown in FIG. 5. Also control oil supply to servomotor 18b in this case is from conduit 134. Conduit 133 feeds and connects to conduit 70 of FIG. 5 which conveys speed governor control oil pressure to pressure transmitter 71. In this case control oil pressure is substituted for first stage exhaust steam pressure. The changes in control air output signal in conduit 72 are proportional to changes in control oil pressure in conduit 133; otherwise the FIG. 5 control system is the same as previously described. Function generator 73 characterizes inlet steam pressure alng the line FU-R on FIGURE 3. Thus the turbine speed governor controls valve 6 up to load R. Above load R, the speed governor controls valves B, C. D and E.

A combination of the two control systems is required to control inlet steam to the high pressure turbine 30 along line FUV-J as shown in FIG. 3. FIG. 7 is a modification of FIG. 5 for this case. The additions on FIGURE 7 are the circuits from conduit 133 to and including the relay 193 in the conduit between items 73 and 75. Conduit 133 (FIG. 6) conducts control oil pressure to pressure transmitter 189 which converts changes in control oil pressure in conduit 133 to proportional changes in control air pressure in conduit 190 in a low pressure range (i.e. 327 p.s.i.g.). Function generator 191 is similar to item 73 except conduit 190 connects to the B chamber in item 191 which modifies the control action formula to: Y(B+X)f(B)=D. Where D is the control air output signal in conduit 192, X and Y are constants, B is the control air pressure in conduit 190, f is a function generated by a cam positioned by control air pressure in conduit 190. Conduit 192 feeds to relay 193 which proportionally subtracts the signal in conduit 192 from the signal in conduit 74. The control action formula of relay 193 is: C-|-Y-(B+X)G:D. Where D is the output control air signal, C is the control air signal in conduit 74, B is the control air signal from conduit 192, Y and X are constants and G is the gain or proportional band setting.

The output signal from relay 193 in conduit 194 feeds to 75 and is the set point for control of steam pressure downstream of valve 6.

The control air signal from function generator 191 in conduit 192 is shaped so that its effect in relay 193 is zero from loads U to V to J as shown in FIG. 3.

Thus, FIG. 7 functions like FIG. 5 in this range. At this time the proportional range (gain) of the turbine speed changer may be adjusted and optimized as described for pivot positioner 162 shown in FIG. 6.

When load U is reached and as load declines further, governor valves A are held open through the action of pressure reducing valve 185 shown in FIG. 6 which maintains a minimum control oil pressure in servomotors 18a and 18b. The declining control oil pressure feeds through conduit 133 (FIGS. 6 and 7) to pressure transmitter 189 (FIG. 7) and function generator 191 increases the control air signal in conduit 192 thus decreasing the set point in conduit 194. In effect function generator 73 shapes the line BB-U shown in FIG. 3 and function generator 191 shapes the differential between lines BBU and F-U shown in FIG. 3. Subtracting the output from function generator 191 from the output of 73 in relay 193 shapes the line FU also shown in FIG. 3.

In the above descriptions of FIGS. 5, 6 and 7 the conrol air circuits only are shown. The power air supplies to all devices and relays has not been shown. One versed in the control art will understand the application of this system.

It will be noted that the above systems are simple and coordinate operation of valve 6 and the turbine speed governor to parallel operation of other electric generating units. The speed changer items 120, 121, 122, 123, 124 and shown in FIG. 6 can be coordinated with other systems for integrating steam generator and turbinegenerator operation or for controlling the turbine-generator electrical output from an external signal associated with economical production of power and frequency control, all in a conventional manner.

Where turbine inlet steam pressure is controlled along line U-V as shown in FIG. 3, to limit maximum first stage blade stress, overpressure protection device may be desired in the event of failure of valve 6 controls. These may consist of the following:

Pressure differential may be measured from conduits 77 to 70 shown in FIG. 5. When the pressure differential exceeds a preselected high value, a switch closes which energizes a relay which closes the motor starter to run the speed changer motor 123 (FIG. 6) in a direction to reduce the compression on spring 119 so as to reduce control oil pressure and close the governor valves. The speed changer motor continues to run to close the gover nor valves until the pressure differential is restored to the predetermined maximum limit.

As a secondary protection, after a period of time without correction of high differential pressure, or in the case where differential pressure exceeds a second high limit, a switch closes which actuates a relay which causes the governor valves or the stop valves 8a and 8b or both to trip closed.

Thus, it will be seen that I have provided efiicient methods, devices and systems for improving turbine steam admission controls which control means is coordinated with steam pressure control within the steam generator especially during the lower portion of the load range; furthermore, I have provided means for controlling steam temperatures within the turbine by control of inlet steam pressure in a way which is compatible with the proper functioning of the turbine speed governor control system and in a way which will improve turbine cycle efficiency, particularly with respect to the reheat portion of the cycle; also, I have provided means for controlling pressure drop across the turbine first steam expansion stage below a preselected limit for reducing first stage blade stress; also, I have provided means for coordinating operation and control of the turbine governor valves with a pressure reducing valve located internally within the steam generator circuits.

While I have illustrated and described several embodiments of my invention, it will be understood that these are by way of illustration only, and that various changes and modifications may be made within the scope of the following claims.

I claim:

1. A steam electric generating plant of the reheat type and under variable load condition including a steam generator having a primary fluid circuit a part of which is composed of steam generating and superheating heat absorption conduits, a power operated throttling valve means located in said primary fluid circuit and intermediately between said heat absorption conduits, a high pressure steam turbine having a partial admission step controlled first steam admission expansion stage including segmented stationary steam admission nozzle groups flowing steam to rotating blades of said first steam admission expansion stage, governor valves for controlling flow of steam to said segmented stationary steam admission nozzles, connecting conduit means from said steam generator primary fluid circuit outlet to said governor valves and from said governor valves selectively to said segmented stationary steam admission nozzle groups, power actuator means for said governor valves adapted to open and close at least a portion of said governor valves sequentially to increase and decrease steam flow from said steam generator to said high pressure turbine, said sequential opening and closing of said governor valves correspondingly increasing and decreasing the number of said nozzle groups flowing steam, control means including said power operated throttling valve means and said power actuator means for said governor valves adapted to selectively open and close said throttling valve means and to simultaneously and selectively position said governor valves to respectively increase and decrease steam flow to said high pressure turbine and to respectively increase and decrease steam pressure in said connecting conduit means upstream of said governor valves throughout at least a lower portion of said variable load condition.

2. A steam-electric generating plant as defined in claim 1 and including means for reheating steam exhausting from said high pressure steam turbine as a part of said steam generator, 21 reheat steam turbine, conduit means for conveying said steam exhausting from said high pressure turbine to said means for reheating steam and for conveying said steam a-fter reheating to said reheat steam turbine, said steam generator primary fluid circuit, said control means and said means for reheating steam together being adapted for increasing steam enthalpy entering said high pressure turbine as said steam pressure in said connecting conduit means upstream of said governor valves is first decreased by said control means and for simultaneously adjusting enthalpy of said steam eX- hausting from said high pressure turbine within the capability of said means for reheating steam to maintain substantially constant steam temperature at said reheat steam turbine inlet.

3. A steam-electric generating plant as defined in claim 1 and wherein said control means is adapted to limit the maximum pressure drop across any portion of said steam admission nozzle groups by reduction of steam pressure in said connecting conduit means upstream of said governor valves by closure of said throttling valve means.

4. A steam-electric generating plant as defined in claim 1 and wherein said control means is responsive to load index means for said steam-electric generating plant.

5. A steam-electric generating plant as defined in claim 4 and wherein said control means is also responsive to said high pressure turbine speed change.

6. A steam-electric generating plant as defined in claim 5 and wherein said selective positioning of said governor valves includes opening and closing respectively with decreases and increases in change of high pressure turbine speed, said control means including proportional control for the selective positioning of said governor valves and means for varying said proportional control. gain responsive to load change.

References Cited by the Examiner UNITED STATES PATENTS 2,184,224 12/1939 Lucke 60-105 2,201,622 5/1940 LaMont 60105 2,811,837 11/ 1957 Eggenbe-rger 60-73 3,069,859 12/ 1962 Weehuizen 6073 SAMUEL LEVINE, Primary Examiner.

ROBERT R. BUNEVICH, Examiner. 

1. A STEAM ELECTRIC GENERATING PLANT OF THE REHEAT TYPE AND UNDER VARIABLE LOAD CONDITION INCLUDING A STEAM GENERATOR HAVING A PRIMARY FLUID CIRCUIT A PART OF WHICH IS COMPOSED OF STEAM GENERATING AND SUPERHEATING HEAT ABSORPTION CONDUITS, A POWER OPERATED THROTTLING VALVE MEANS LOCATED IN SAID PRIMARY FLUID CIRCUIT AND INTERMEDIATELY BETWEEN SAID HEAT ABSORPTION CONDUITS, A HIGH PRESSURE STEAM TURBINE HAVING A PARTIAL ADMISSION STEP CONTROLLED FIRST STEAM ADMISSION EXPANSION STAGE INCLUDING SEGMENTED STATIONARY STEAM ADMISSION NOZZLE GROUPS FLOWING STEAM TO ROTATING BLADES OF SAID FIRST STEAM ADMISSION EXPANSION STAGE, GOVERNOR, VALVES FOR CONTROLLING FLOW OF STEAM TON SAID SEGMENTED STATIONARY STEAM ADMISSION NOZZLES, CONNECTING CONDUIT MEANS FROM SAID STEAM GENERATOR PRIMARY FLUID CIRCUIT OUTLET TO SAID GOVERNOR VALVES AND FROM SAID GOVERNOR VALVES SELECTIVELY TO SAID SEGMENTED STATIONARY STEAM ADMISSION NOZZLE GROUPS, POWER ACTUATOR MEANS FOR SAID GOVERNOR VALVES ADAPTED TO OPEN AND CLOSE AT LEAST A PORTION OF SAID GOVERNOR VALVES SEQUENTIALLY TO INCREASE 